Method for regulating a clutch or a brake in a transmission

ABSTRACT

The invention relates to a method for regulating an electrohydraulically controlled clutch or a brake of a transmission. According to the invention, the clutch or brake is regulated using a model-based compensation pressure regulator with the help of an observation unit, whereby the compensation-pressure control circuit contains a non-linear compensation element which corresponds to the inverse model of the control system of the coupling or brake. The observation unit estimates interference levels of the clutch regulation from a drive chain model, according to a condition-estimation procedure.

FIELD OF THE INVENTION

The invention relates to a method for regulating a clutch or a brake inan electrohydraulically controlled transmission.

BACKGROUND OF THE INVENTION

An automatic transmission for a motor vehicle usually has severalclutches and brakes with which different gear steps and the direction oftravel can be shifted. In such automatic transmissions there iscustomarily used as starting element a hydrodynamic torque converterwhich, to optimize the total efficiency of the transmission, is providedwith a converter lock-up clutch which in certain gear steps is partly ortotally closed depending on rotational speed and load of the drivenengine.

There are known also special wet operating starting clutches which canalso be integrated in the automatic transmission. But those startingclutches are mostly situated as unit on the transmission input, for ex.,also as starting element in an automatic continuously variabletransmission. The wet starting clutch is also disposed as unit on thetransmission output, specially in CVT transmissions (continuouslyvariable transmissions).

The clutches and brakes in the automatic transmission or in theautomated continuously variable transmission are usuallyelectrohydraulically controlled independently of the application of theshifting element as gear-change clutch, starting clutch, or converterlock-up clutch. For this purpose variables according to specifics of thevehicle, specifics of the transmission and driving mode such as axle andgear ratios, transmission oil temperature, engine torque, enginerotational speed, vehicle speed, internal rotational speeds of thetransmission, accelerator pedal position and accelerator pedal change,brake signal, vehicle acceleration, tractional resistance and driveractivity are processed in an electronic transmission control unit andcorresponding output signals relevant to the clutch, for ex., forpressure control or pressure regulation, or during a gear change or forrotational speed regulation during a slip operation, are transmitted toa hydraulic control unit and there converted by means of actuators andhydraulic valves into hydraulic pressure for the corresponding clutch.

Gruhle, Jauch, Knapp and Ruchardt describe as effective example in VDIReports No. 1175, 1995, a method for model-supported application of aregulated converter lock-up clutch in automatic transmissions ofpassenger cars. The converter lock-up clutch and the control thereof areat present a sub-system in the automatic transmission which decisivelyimprint the driveability of a vehicle. To meet the increasingrequirements in comfort, driving performance and fuel consumption, hereis proposed a complex control draft of the converter clutch providedwith a control loop which works with the differential rotational speedfrom engine and turbine rotational speeds. To stabilize and increase thedynamics of the control loop, there are implemented in addition a loadmodulation of the engine torque and special functions with directpressure standards for the converter lock-up clutch.

In its basic structure the control draft described by way of example isapplicable to all electrohydraulically actuated clutches and brakes intransmissions of passenger cars and is not confined to the case of theconverter lock-up clutch. But disadvantages in such a complex controldraft are the basic structure of the software of the electronictransmission control which is vague due to added functionalities and canbe extended only at great expense, and the considerable cost forapplication of such functions.

Therefore, the problem on which this invention is based is, departingfrom the cited prior art, further to develop with regard to improvedcontrol quality and control dynamics and reduced application cost, amethod for regulating a clutch or brake designed for ex., as gearclutch, starting clutch or converter lock-up clutch.

According to the invention this problem is solved by the features ofclaim 1. Other developments of the invention result from the sub-claims.

SUMMARY OF THE INVENTION

The invention accordingly proposes to design the clutch regulation witha model-based compensation pressure regulator using an observation unitof interference level which, based on a drive train model, estimatesinterference levels of the clutch regulation according to acondition-estimation method. The interference levels reproduce theinaccuracies in comparison with the real system of the physicalmathematical model used according to regulation technology and result,specially from characteristic line errors determined by principle, instationary and dynamic control errors, hydraulic tolerances, the same asdynamic model errors determined by principle.

Therefore, the observation unit estimates from the drive train model theinterference torque on the clutch or brake to be shifted or regulatedand the load torque cropping up in power flow direction behind theclutch or brake. According to the configuration of the drive trainmodel, the characteristic shape of the torsional oscillations of thedrive train can also be estimated.

The inventive method is based on a non-linear compensation methodaccording to the principle of exact linearization. From the linkage of ahydraulics model of the clutch control with an inverse model of thecontrol system, a simple transmission function results. The remainingsystem can be drafted by control technology according to the lineartheory specially by a separation of the non-linear portion of theobservation unit.

By such a regulation draft of the clutch constructed as compensationpressure control loop supported by observation unit, there are obtainedspecially advantageously an operating-point dependent, uniform controlbehavior of the closed control loop with high adjusting dynamicsresulting therefrom and a strong sequence and interference behavior.

In a first development of the invention it is proposed to design theinterference-level observation unit as reduced observation unit whichestimates only the unknown system variables relevant for the clutchregulation. Hereby is advantageously obtained as high as possible acomputer speed and therewith as high as possible regulator dynamics.

In a second development of the invention it is proposed to design theinterference-level observation unit as complete observation unit whichestimates all system variables. Hereby can be advantageously improved inparticular a servo-control quality for the regulator and therewith thequality of the regulator.

In a third development of the invention it is proposed to design theinterference-level observation unit as Kalman filter which isadvantageously sturdy in relation to signal rustles.

In another development of the invention it is proposed to design asseveral linked regulation blocks the regulator of the closed,model-based and observation-unit supported compensation pressure controlloop. A first block processes as rotational speed regulator therotational speeds and slip standards of the clutch regulation. Therotational speed regulator advantageously does not need to containcomponents for ensuring stationary precisions and can be designed assimple P regulator. A consecutive second block calculates from thetheoretical standards of the rotational speed regulations, from theservo-control values of the inverse system model and from estimatedinterference levels of the observation unit, a pressure standard for thehydraulic servo component of the clutch control. The conversionindependent of the operating point of the regulated quantity isadvantageous here. A consecutive third block compensates as pressureregulator the tolerances of the hydraulic servo component with the aidof estimated variables of the observation unit and delivers atheoretical pressure for the clutch control. By the compensation of thetechnical pressure fluctuations of customary pressure adjusters in thewhole operating range, the dynamics of the control loop is clearlyimproved. A consecutive fourth block works as traditionalpressure-current regulator and converts the theoretical pressurestandard into a current standard for the electrohydraulic pressureadjuster.

In one other development of the invention it is proposed to design therotational speed regulator of the closed, model-based andobservation-unit supported compensation pressure control loop as PDregulator with additional non-linear term. In case of great divergencesfrom a theoretical value, a quicker approximation to the theoreticalvalue is hereby achieved.

In another development of the invention it is proposed to design thepressure regulator of the closed, model-based and observation-unitsupported compensation pressure control loop as regulator with PT2characteristic according to the principle of exact linearization. Thedynamics of the control loop can be advantageously influenced hereby.

The inventive regulation draft at the same time can be universally usedfor every electrohydraulically controlled clutch. Cases of applicationare, for ex., a wet starting clutch in an automatic transmission or inan automated manual transmission without or with stand-by control SBC, aselector clutch or selector brake in an automatic transmission or aconverter lock-up clutch in an automated transmission.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention is explained in detail herebelow with reference todrawings based on the example of a converter lock-up clutch. In thedrawing:

FIG. 1 is a diagrammatic representation of a simplified hydrauliccontrol of a converter lock-up clutch,

FIG. 2 shows by way of example identification by measure technology ofthe model equation for the non-linear transmission behaviorp_WK=f(i_EDS),

FIG. 3 is a diagrammatic representation of three drive train models asFIG. 3-1, FIG. 3-2 and FIG. 3-3,

FIG. 4 shows by way of example an inventive compensation pressurecontrol loop with an observation unit of interference level and

FIG. 5 shows a differential equation system for an observation unit ofinterference level designed by way of example.

DETAILED DESCRIPTION OF THE INVENTION

As fundamental in every clutch, the pressure differential on the pistonof the converter lock-up clutch (WK) determines the torque transmissibleby the clutch. Assuming that the pressure portions generated by therotation are of equal magnitude on both sides of the clutch, the supplypressure of the WK piston determines alone the torque transmissible withthe WK

 M _(—) WK=A_piston*r_piston*z*μ*p _(—) WK

with

A_piston surface of the WK piston R_piston central friction liningradius z number of friction surfaces μ friction value = f (temperature,differential rotational speed, lining compression) p_WK supply pressureto the WK piston.

The geometric variables and the friction value can be combined in onevariable μ_geo there resulting for the torque transmissible with the WK:

M _(—) WK=μ_geo* p _(—) WK.

Diverging from the ideal behavior of the clutch, the following effectscan now be observed:

In case of differential rotational speeds unequal to zero, torque istransmitted parallel to the WK via the hydraulic control loop of theconverter. The current in the loop also affects the space between theturbine wheel and the WK piston.

The friction linings are grooved to cool the friction surfaces. The oilflowing through the grooves leads a large part of the generated heatdirectly away from the generation point. The radial current betweenpiston and lid forms a more or less strongly defined whirl which changesthe pressure between housing and WK piston.

Depending on the length and cross section of the recoil pipe and of theflow rate, a back pressure generates. When said effects are taken intoaccount, the torque equation reads:

M _(—) WK=μ_geo*(p _(—) WK−p_0)

 with

p_0 pressure offset=f (engine rotational speed, through flow).

μ and μ_geo respectively and p_0 cannot be calculated with enoughaccuracy. The values are determined by measurements and integrated intocharacteristic fields in the model. The torque absorption and deliveryof the torque converter are taken into consideration in the model withtwo characteristic lines which describe the stationary behavior asfunction of the rotational speed ratio of turbine rotational speed toengine rotational speed:

M_p2000 = f(v) pump absorption torque at engine rotational spend n_mot =2000 1/min μ_WV = f(v) converter reinforcement ν = n_t/n_mot rotationalspeed ratio of turbine rotational speed to engine rotational speed

With said characteristic lines it is possible for each operating stateto calculate the pump torque M_P and the turbine torque M_T of thetorque converter to

M _(—) p=M _(—) p2000(ν)*n_mot**2/(2000 l/min)**2 and

M _(—) t=μ _(—) WV(ν)*M _(—) p.

FIG. 1 shows now a diagrammatic representation of an extensivelysimplified hydraulic control of a converter lock-up clutch. With MKV isdesignated a converter clutch valve through which the piston of theconverter lock-up clutch WK is loaded for torque transmission withpressure p_WK. The WK usually has a certain oil through flow Q_WK, forex., for cooling the lining in the permanent slip state. This WK throughflow is shown in FIG. 1 simplified by a diaphragm. A system pressurep_HD is first reduced via constant pressure regulating valve DRV to apressure le l tuned to the type of the electric pressure adjuster EDS,for ex., an electrohydraulic pressure regulator or an electrohydrauliccycle valve. The EDS is loaded by the electronic transmission controlwith a control current i_EDS whereby the converter clutch valve WKV,which is designed, for ex., an analog valve, is loaded with a controlpressure and releases the pressure line from the system pressure p_HD tothe piston of the converter lock-up clutch. The converter clutch valveWKV adjusts here a pressure p_WK tuned to the converter type and to thetorque to be transmitted. The hydraulic elasticities c_hydr are plottedfor further clarification of the influences acting upon the controlsystem.

The non-linear stationary transmission behavior of the converter clutchpressure p_WK=f (i_EDS) produced by the pressure adjuster EDS is shownas characteristic line according to control technology in the model. Thedynamic simulation of the system generates a non-linear behavior of2^(nd) order which according to the analysis of the frequency responsescan be described with the following model equation

T(0, p _(—) WK)**2*p _(—) WK+2*D(0, p _(—) WK)*T(0, p _(—) WK)p _(—)WK+p _(—) =p_soll

with

T(0,p_K) time constant as characteristic field according to transmissiontemperature θ and pressure p_WK, D(0,p_(wk)) damping as characteristicfield according to transmission temperature θ and pressure p_WK.

FIG. 2 shows by way of example an identification according to measuringtechnology of the model equation for the transmission behaviorp_WK=f(i_EDS). A variation is shown of the time constant and lossesdepending on the temperature and on the pressure with reference to thefrequency responses and phase responses.

The mathematical description of the non-linear transmission behavior ofthe converter can obviously be improved by detailed physicalmathematical models which can simplify the identification by testtechnology of the individual equation parameters.

In FIG. 3 are reproduced by way of example three simplifiedtorque-volume models of the drive train of the vehicle. On account ofthe changing inertia torques in the transmission, the models aredescribed according to gear and gear shift. FIG. 3-1 shows a simplemodel for a fixed gear step that can be described with two differentialequations:

J_mot*{dot over (ω)}_mot=M_mot−M _(—) p−M _(—) WK

J _(—) G*{dot over (ω)} _(—) t=M _(—) t+M _(—) WK−M _(—) L

with

M_mot torque produced by the engine reduced by all losses (friction,added auxiliary aggregates including transmission pump), M_p, M_t pumpand turbine torque of the hydraulic loop of the converter, M_WK torqueon the converter lock-up clutch, M_L load torque from tractionalresistance on the output, J_mot engine inertia torque (including primaryside of the converter), J_G gear-dependent substitute inertia torque(contains secondary side of the converter, all rotating parts of thetransmission, universal shaft, differential gear, axle shafts, wheelsand vehicle mass; respectively reduced to the turbine rotational speedn_t), ω_mot angular velocity of the engine, ω_t angular velocity of theconverter turbine.

In FIG. 3-2 is shown a second model for a drive train which takes intoaccount a gear shift, that is, quitting the synchronous rotational speedof the old gear till reaching the synchronous rotational speed of thenew gear. The model can be described by the equations:

J_mot*{dot over (ω)}_mot=M_mot−M _(—) p−M _(—) WK

J _(—) G*{dot over (ω)} _(—) t=M _(—) t+M _(—) WK−M _(—) K+(k _(—) L*M_(—) L)

M_mot torque produced by engine, reduced by all losses (friction, addedauxiliary aggregates including transmission pump), M - p, M_t pump andturbine torque of the hydraulic loop of the converter, M_WK torque onthe converter lock-up clutch, J_mot engine inertia torque (includingprimary side of the J_G converter), shift-dependent inertia torque ofthe transmission (containing all rotatory parts between WK and gearclutch; reduced to the turbine rotational speed n_t), M_K torque on thegear clutch, (k_L*M_L) portion of the load torque from kinematiccoupling via the planetary gear set, ω_mot angular velocity of theengine, ω_t angular velocity of the converter turbine.

Simulation and tests in the vehicle have shown that the simple drivetrain model of FIG. 3-1 is no longer adequate for large total ratios(that is, for low gears), since it disregards the first vibrationcharacteristic form of the drive train. The spring rigidities of theaxle shafts are determinant here. In FIG. 3-3 is shown by way of examplea model that takes into account the spring rigidities of torsionalvibrations and damping of torsional vibrations of the drive train. Thenumber of differential equations increases here to three:

J_mot*{dot over (ω)}_mot=M_mot−M _(—) p−M _(—) WK

J _(—) G*{dot over (ω)} _(—) t=M _(—) t+M _(—) WK−M_spring

J _(—) F*{dot over (ω)} _(—) ab=M_spring−M_(—) L

with

M_spring=c*(∫ω_(—) tdt−∫ω _(—) abdt)+d*(ω_(—) t−ω _(—) ab)

and

M_mot torque produced by the engine reduced by all losses (fricition,added auxiliary aggregates including transmission pump), M_p, M_t pumpand turbine torque of the hydraulic loop of the converter, M_WK torquecorresponding to the deflection and deflection velocity of the axleshafts, M_spring torque corresponding to the deflection and deflectionvelocity of the axle shafts, M_L load torque from tractional resistanceon the output, J_mot inertia torque of the engine (including primaryside of the converter), J_G gear-dependent inertia torque oftransmission and output drive train (contains all rotating parts betweenWK and shift clutch J_F axle shafts, reduced to the turbine rotationalspeed n_t), gear-dependent inertia torque of the vehicle (containswheels and vehicle mass reduced to turbine rotational speed n_t) ω_motangular velocity of the engine, ω_t angular velocity of the converterturbine ω_ab angular velocity of the output.

Other more complex mathematical drive train models can obviously beshown, for ex., the model of FIG. 3-2 can be extended by the influencesof the torsional elasticities similarly to the arrangement of the modelof FIG. 3-3.

According to the invention, with the aid of the system informationdescribed above by way of example, a model-based regulation is nowdevised which makes available the desired functionality without it beingnecessary in certain operating situations to refer to special functions.High dynamics of the closed control loop are implemented, the regulationdivergences remaining therefore sufficiently small in all cases duringhigh dynamic procedures. Besides, the regulation is here so strong thatthe control quality, despite the technically unavoidable scattering ofparts of all components contained in the control loop and the systemchanges occurring in the course of the transmission life, is alwaysmaintained.

The requirement of greater strength represents in a certain manner acontradiction to the requirement of high dynamics, since great circuitreinforcements generally lead to a reduction of the stability reserves.The contradiction according to the invention is solved by a precisecompensation and servo-control of all known influences and the regulatorthus becomes extensively released. The other advantages of the inventiveregulation are the easier applicability by parameters mostly physicallyinterpretable and the reduction of the parameters by about 50% comparedto the known function.

FIG. 4 shows a designed example of an inventive control system forclutch regulation with model-based compensation pressure control loopwith the aid of an observation unit of interference level which from thedrive train model estimates the interference levels of the clutchregulation according to a condition estimation method. The block diagramof the control loop consists of the blocks “signal preparation”,“estimated value generation”, “inverse system model”, “observationunit”, “rotational speed regulator”, “torque-pressure conversion”,“pressure regulator”, “pressure-current conversion” and “controlsystem”. Both the “inverse system model” and the “observation unit”underlie here the above described model equations.

The “inverse system model” compensates, with the aid of the measuredrotational speeds and of the engine torque made available by theelectronic engine control of the drive engine or detected by measuretechnology, all known torques acting upon the system. A correspondingservo-control torque M_vor can be described by the equation${M\_ vor} = {{\frac{J\_ d}{J\_ mot} \star {M\_ mot}} - {\frac{J\_ d}{J\_ mot} \star {M\_ p}} - {\frac{J\_ d}{J\_ G} \star {M\_ t}} - {J \star {{\omega\_ d}{\_ soll}}}}$

The rotary torques are now advantageously related to the differentialangular velocity of the converter lock-up clutch to be regulated. J_D ishere the inertia torque resulting from the differential angularvelocity. For a full compensation are also needed the load torque M_Labutting on the output and determined by tractional resistance, the sameas in gear shifts the torques of the clutch to be engaged anddisengaged. Said torques can be combined in one total shift torque M_K.If load torque M_L and total shift torque M)K are in addition modulatedin adequate manner, the system is in a balanced state of stable limitindependently of the actual operating point. But this statement appliesonly to stationary operating states due to the delays introduced by thehydraulic system.

Since the physical mathematical system model cannot fully reproduce thebehavior of the real complex system, according to the invention there isintroduced in the model equations an interference torque M_S where arecombined the differences between model and reality. On account of therelatively great uncertainties of the parameters in the converter clutchobserved by way of example, said interference is taken into account asadded clutch torque in the compensation control loop presented by way ofexample. But his constitutes no limitation for a full compensation,since therewith can always be estimated in connection with the likewiseunknown load or shift torques a combination of torques which can coverall conceivable divergences.

Particularly in order to reduce the estimation errors in dynamicprocedures, all the pressure acting upon the clutch piston, that is, thenominal friction torques of the clutch, must be known. Since theygenerally are not shown, there is used according to the invention acombined observation unit of state and interference which reconstructsthe clutch pressure p_WK, load torque M_L and interference torque M_S onthe basis of the rotational speeds measured. During gear shifts thetorques of the shift clutch M_K are observed instead of the load torqueM_L. In the example shown, the load torque M)_L is assumed as constantfor the time period of the gear shift, in another embodiment of thecompensation control loop a variable load moment can also be processedaccording to control technology.

In order to keep low the costs and at the same time increase theperformance capacity of the estimation method, it is proposed accordingto the invention to estimate by observation unit only the unknownvariables, that is, to refer to the draft of a reduced observation unit.The differential equation system for an observation unit thus designedfor estimating the interference levels during a gear shift according tothe drive train model of FIG. 3-2 is indicated as matrix in FIG. 5. Thedifferential equation system of the reduced observation unit cantherefore be represented as vector equation in the form

{dot over (x)}=A*x+B*u.

Outside the gear shifts, the shift torque M_K is replaced in the statevector by the load torque M_L which then drops out from the vector ofthe regulated quantity. When the system matrix A shown in FIG. 5 isdivided in four parts A11, A12=0, A21 and A22, the same as when the setmatrix B shown in FIG. 5 is correspondingly divided in two parts B1 andB2, there results the mathematical form needed for a reduced observationunit draft. The matrix A11 comprises here the square sub-matrix byvirtue of which the variables to be observed retro act upon themselves.The assumptions M_S=const and M_K=const apply to stationary operatingconditions at least for the unknown torques in the dynamic model. Thedifferential equation system of the reduced observation unit has thusthe simplified mathematical form

Herein y=((ω_dω_t)**T describes the vector of the measured parameterswhile from the vector of the measured variables and from the equation

p=(M _(—) S M _(—) K dp _(—) WK/dt p _(—) WK)**T−L*y

the variables to be observed can be reconstructed with the aid of theparameters measured. The matrix L of the observation unit determines thecharacteristic values of the observation unit. With only two free draftparameters 11 and 12, for ex., an observation unit matrix$L = \begin{pmatrix}{\frac{{{J\_ G} \star J} - d}{{J\_ G} + {J\_ d}} \star {l1}} & {\frac{{J\_ G} \star {J\_ d}}{{J\_ G} + J - d} \star {l1}} \\{{- {J\_ mot}} \star {l2}} & {- \left( {{J\_ mot} + {{J\_ G} \star 12}} \right)} \\0 & 0 \\0 & 0\end{pmatrix}$

can be selected for the state estimation. This form advantageouslyresults in that the estimated clutch pressure p_WK is independent ofmeasurements depending only on the regulated quantity p_soll. It couldbe demonstrated that this sub-system of the compensation pressurecontrol loop it itself stable according to control technology andtherefore errors of estimation die away with the characteristic dynamicsof the hydraulic system.

The advantages of this inventive generation of the interference-levelobservation unit are a clear reduction of the computer expenses and abetter interpretability of the total system behavior of the clutchobserved.

Another essential advantage results from the separation of thenon-linear portion of the observation unit, namely, the time constantT=f(θ,p_WK) dependent on the transmission oil temperature θ and on theclutch pressure p_WK and the damping D=f(θ,p_WK) likewise dependent onthe transmission oil temperature θ and on the clutch pressure p_WK. Thedraft of the rest of the system according to control technology can thusresult according to the linear theory.

The limitation to only two draft parameters 11 and 12 in the matrix ofthe observation unit leads to an easier manipulation in the applicationin the vehicle. In the above indicated form said two parameters directlycorrespond to the inverse values of the time constants of the remaininglinear system. Besides, the system, in the closed control loop, isstrong with regard to errors in the hydraulics model when the estimatedpressure does not act directly upon the shift torque which physicallyconditioned possesses high dynamics and hence has to be quickerdetermined then the remaining unknown variables.

In a development of the invention it is proposed to design theinterference-level observation unit so that all system variables beestimated. Hereby can advantageously be further improved in particularthe servo-control quality of the regulator and thus ultimately thecontrol quality.

In another development of the invention it is proposed to design theinterference-level observation unit as Kalman filter, the advantages ofwhich have been mentioned above.

By the formerly described control parts “inverse system model” and“observation unit” of the inventively designed compensationpressure-regulating control loop according to FIG. 4, the system isconverted to a form in which each operating point can be used asstationary balance point according to control technology. This alsonaturally applies to the theoretic value. The function of the block“rotational speed regulator” shown in FIG. 4 now consists in that incase of divergence from the theoretic value, the differential rotationalspeed again approximates it. But because of the above describedinventive steps, this rotational speed regulator does not have tocontain special components to ensure a stationary precision. In thesimplest case the rotational speed regulator can therefore be designedas P regulator. Advantageous for increasing the dynamics is asubstantially more complex PD regulator having an added non-linear termwhich in case of gear divergences makes possible a quicker approximationto the theoretic value and at the same time influences only slightly thebehavior in the proximity of the estimated value. In a PD rotationalspeed regulator thus designed, the non-linear control model finds itsreproduction in the control law for the control torque M_r of the clutchto be regulated

M _(—) r=k _(—) P*(Δω_(—) d+k _(—) NL*|Δω _(d)|*Δω_(—) d)+k _(—) D*Δ{dotover (ω)} _(—) d

with

Δω_(—) d=ω _(—d)_soll−ω_(—) d,

the proportional regulating parameters k_P, the non-linear regulatorparameter k_NL and the differential regulating parameter k_D.

The individual parameters of the rotational speed regulator can becalculated, for ex., by the known method of Gain scheduling. To this endthe system is linearize by the theoretic value, taking into account theother control parts, according to the system of parameters and statevariables, a set of control parameters being determined for eachoperating point so that the characteristic values of the closed controlloop be approximately independent of the actual state of the system. Inorder to reduce here dependencies on other variables of the regulatorcoefficients, it is proposed in the inventive embodiment shown in FIG. 4to introduce in the control loop the blocks “torque-pressureconversion”, “pressure regulator” and “pressure-current conversion”.

The block “torque-pressure conversion” calculates form the estimatedtorque(M_r+M_S+J_d/J_G*(M_K−k_L*M_L)) a theoretic pressure P_nenn forthe converter clutch. In this conversion enter in the first place thefriction value of the clutch lining and a pressure offset which bothdepend mainly on the transmission oil temperature and the enginerotational speed. Hereby is ensured a conversion of the regulatedquantity dependent on operating point.

The sub-system that follows in the block diagram is the “pressureregulator”. An essential importance is attached to this since the systembehavior is decisively influenced by the dynamics of the hydraulicadjusting system. As described above, the dynamics of the adjustingelement essentially depends to the pressure level itself and on thetemperature of the transmission oi. The actually existing fluctuationsof customary adjusting elements are so great in quantity that theycannot be disregarded (usual angular frequencies vary from far above 0.1s to 0.02 s ). Specially in cold operating conditions, the dynamics ofthe adjusting system itself is too low to meet the requirements set onthe regulation with regard to speed. Therefore, the function of theblock pressure regulator is to balance the tolerances of the hydraulicsadjusting system with the aid of the clutch pressure estimated by theobservation unit. The invention proposes to this end a regulator withPT2 characteristic drafted according to the principle of exactlinearization with a control law:${p\_ soll} = {{\left( \frac{T}{T\_ dr} \right)^{2}\left( {{p\_ nenn} - {p\_ WK}} \right)} + {p\_ WK} + {2 \star {\frac{T}{T\_ dr}\left( {{D \star {T\_ dr}} - {{D\_ dr} \star T}} \right)} \star {\overset{.}{p}{\_ WK}}}}$

The regulator parameters T_dr and D_dr are here the time constant andthe damping of the linear PT2 behavior which in principle can bearbitrarily preset as theoretic dynamics. In the practical design limitsare set to high theoretic dynamics by a too strongly differentiatingregulator character. This become important mainly in the above describedstructure of the matrix L o the observation unit. In this dorm thedrafted pressure regulator works as model-supported feed-forwardcompensation of the non-linear pressure dynamics which replaces this bythe desired behavior. The effectiveness of said step evidently dependsdecisively on the precision of the underlying model. In the practice hasbeen found that modeling of the time behavior dependent on operatingpoint is already enough to be able to achieve considerable performancegains compared to the non-compensated system.

The block pressure regulator is followed by a block “pressure-currentconversion” which essentially consists of an inversion of the stationarypressure regulating line with an attached current regulation. Thecurrent regulation is here a standard regulation such as used in commonpressure regulators. A modification is not needed, since the limitingfrequency of the usual current regulation is clearly above that of thepressure regulator.

The last two blocks in the block diagram of the compensation pressureregulation control loop shown in FIG. 4 are the “signal preparation” andthe “theoretic value generation”. Both adequate supply the regulatorwith the needed information. The latter plays an important part chieflyin the “signal preparation”. Besides an eventual filtration of stronglynoisy signals, this block takes care that all information be availablesynchronized in time. The observation unit can specially behavesensitively to a blending of signals picked up synchronically relativeto each other. This results, for ex., from the different CAN-bus runningperiods of individual signals and the preparation thereof. Individualsignals (such as engine rotational speed) also cannot actually be madeavailable in all operating states for each regulation cycle, since theyare fed by an external control unit of the electronic transmissioncontrol. The invention proposes that the synchronization be carried outby delaying quick signals and extrapolating slow signals. Th e“theoreticvalue generation” makes available, together with the theoretic valuew_d_soll for the rotational speed regulation of the clutch, the timedderivation thereof which is advantageous for a correct servo-controlsystem. Said variables are determined, for ex., by using filters ofsecond order which constantly differentiatably readjusting the actualvariables by an ad3quate filter initialization.

The inventive model-based compensation pressure control loop with anobservation unit of interference levels, which estimates from the drivetrain model the interference levels of the clutch regulation accordingto a state estimation method, can be universally used for everyelectrohydraulically controlled clutch. Other cases of applicationbesides the regulation described in detail of a converter lock-up clutchare, for ex., a wet starting clutch in an automatic transmission or anautomated manual transmission without or with stand coupling (stand bycontrol, SBC), a shifting clutch or shifting brake in an automatictransmission. Compared to the control drafts known from the prior art,by an operating-point dependent, uniform regulation behavior of theclosed control loop, essential advantages are obtained relative tosequential and interference behavior and dynamics. These are obtainedby:

compensation pressure regulator to accelerate the servo component,

substitution of an observation unit of interference level for the usualI regulator,

improved servo-control (more accurate model).

The gain in performance capacity makes possible to a great extent toomit special functions, which makes easier both the manipulation and thestructuration o the software. The model-based control draft makespossible the use of physically interpretable parameters. Thisconsiderably simplifies the [German word “Portierbarkeit” forth coming]and application.

References WK converter lock-up clutch WKV converter clutch valve p_WKconverter clutch pressure p_HD system pressure DRV constant pressureregulating valve EDS electrohydraulic pressure regulator i_EDS currentQ_WK WK through flow c_hydr hydraulic elasticity M_mot engine torqueJ_mot inertia torque of engine and primary side of converter M_WK torqueon the converter lock-up clutch J_G gear-dependent substitute inertiatorque of transmission and drive train M_L load torque of tractionalresistance M_K torque on a shifting clutch d torsional vibration-dampingconstant of the drive train c torsional vibration-spring rigidity of thedrive train J_F gear-dependent inertia torque of the vehicle M_r outputvalue of torques of the rotational speed regulator M_vor servo-controltorque of the inverse system model M_K estimated clutch torque k_L * M_Lportion of the load torque from kinematic coupling via the transmissionwheel sets J_d inertia torque generating from the differential angularvelocity of the clutch to be regulated ω_mot angular velocity of theengine ω_t angular velocity of the converter turbine ω_d differentialangular velocity on the clutch to be regulated ω_d_soll estimated valueof the differential angular velocity on the clutch to be regulatedp_nenn pressure output value of the torque-pressure conversion p_sollpressure output value of the pressure regulator A system matrix of thedifferential equalization of a reduced observation unit B adjustingmatrix of the differential equalization of a reduced observation unit Ttime constant D damping constant

What is claimed is:
 1. A method for regulating an electrohydraulicallycontrolled clutch and brake with a control system for a clutch or brakeof a transmission, the method comprising the steps of: regulating theclutch or brake using a model-based compensation pressure regulator withthe aid of an observation unit; providing an electronic control loop ofsaid compensation pressure regulator containing a non-linearcompensation value which corresponds to an inverse model of the controlsystem for the clutch or brake; and estimating using said observationunit, interference levels of the clutch regulation on the basis of adrive train model according to a condition-estimation procedure.
 2. Themethod according to claim 1, further comprising the step of using saidobservation unit to estimate only the unknown system variables relevantfor the clutch regulation.
 3. The method according to claim 1, furthercomprising the step of using said observation unit to estimate allsystem variables relevant for the clutch regulation.
 4. The methodaccording to claim 1, further comprising the step of designing saidobservation unit as Kalman filter.
 5. The method according to claim 1,wherein said compensation pressure control loop contains a regulatorwhich from measured signals, theoretical standards of the measuredsignals, calculated values of the inverse system model of the clutch orbrake and from at least one estimated value of the observation unit,determines an operating-point independent regulated quantity.
 6. Themethod according to claim 1, wherein the compensation pressure controlloop contains a regulator which from measured signals, theoreticalstandards of the measured signals and from at least one estimated valueof the observation unit, forms a regulated quantity independent of thetolerance of an electrohydraulic servo component through which theclutch of brake is hydraulically controlled.
 7. The method according toclaim 1, wherein said compensation pressure control loop contains aregulator which from a calculated pressure standard for anelectrohydraulic servo component through which the clutch or brake ishydraulically controlled and from at least one estimated value of theobservation unit, determines a regulated quantity independent of thetolerance of the electrohydraulic servo component.
 8. The methodaccording to claim 1, comprising the step of using the compensationpressure control loop contains several linked regulation blocks whereina first regulator block (rotational speed regulator) to process measuredrotational speeds and rotation speed estimated values or the equivalentsignals thereof (ω_t,ω_mot,ω_d,ω_d_soll) and forming, using a secondconsecutive regulator block (torque-pressure conversion) from thetheoretical standard (M_r) of the first regulator block, fromservo-control vales (M_vor) of the inverse system model (M_S,J_d/J_G*(M_K−k_L*M_L)) of said observation unit, an operating-pointindependent regulated quantity (p_nenn), and forming, using a thirdconsecutive regulator block (pressure regulator) from the theoreticstandard (p_nenn) of said second regulator block and from the estimatedinterference levels (p_WK) of said observation unit, a regulatedquantity (p_soll) independent of the tolerance of the electrohydraulicpressure adjuster and forming, using a fourth consecutive regulatorblock (pressure-current conversion) forms from the theoretic standard(p_soll) of said third regulation block a current standard (i_EDS) withwhich the electrohydraulic pressure adjuster is loaded.
 9. The methodaccording to claim 8, further comprising the step of designing saidfirst regulator block (rotational speed regulator) as a P regulator. 10.The method according to claim 8, further comprising the step ofdesigning said first regulator block (rotational speed regulator) as PDregulator with added non-linear term.
 11. The method according to claim8, wherein said third regulator block (pressure regulator) is designedas regulator with PT2 characteristic according to the principle of exactlinearization.
 12. The method according to claim 1, wherein a timedderivation of theoretical values of the rotational speed of the clutchregulation or equivalent variables thereof are processed in the inversemodel of the control system.
 13. A method for regulating anelectrohydraulically controlled clutch and brake with a control systemfor a clutch or brake of a transmission, the method comprising the stepsof: regulating the clutch or brake using a model-based compensationpressure regulator with the aid of an observation unit; providing anelectronic control loop of said compensation pressure regulatorcontaining a non-linear compensation signal which corresponds to aninverse model of the control system for the clutch or brake; estimatingusing said observation unit interference levels of the clutch regulationon the basis of a drive train model according to a condition-estimationprocedure; and providing in said electronic control loop a controllerthat generates a working point independent control variable on the basisof measured signals, setpoint assignments of the measured signals,calculated values of the inverse model of the control system of theclutch or brake, ad at least one estimated value of the observationunit.